EP1525394B1 - Joint arranged between mutually mobile parts of a hydraulic machine - Google Patents

Description

The subject invention relates to a device for sealing a gap between two relatively movable parts of a hydraulic machine with at least one sealing body, which is mounted relative to the two movable parts by means of at least one hydrostatic bearing, each of the hydrostatic bearing consists of mutually facing bearing surfaces and wherein at least one bearing surface at least one bearing element, such as a groove, groove od. Like. Has, which is supplied via at least one supply line with a hydraulic storage medium. Furthermore, the invention discloses a method for operating and a method for starting up such a seal.

Due to the very unfavorable operating conditions in the range of the outer diameter of an impeller of a hydraulic machine, and especially for larger machines with impeller diameters of up to several meters, it has not been possible to build a reliable seal between impeller and housing, which in addition to a loss of efficiency By gap losses can also lead to other significant problems. This is mainly due to the fact that in the range of the outer diameter of the impeller, very high peripheral speeds occur, the impeller, and the housing are subjected to strong vibrations and due to the high-acting pressures, the impeller learns additional axial displacements. These operating conditions have hitherto prevented the construction of a nearly, or ideally complete, tight seal. Previously applied seals, such as e.g. Labyrinth seals, are not seals in the strict sense, but only devices to reduce the fissile water flow. Others, e.g. Known ice ring seals, in turn, were very expensive and unreliable.

In the axial wheel side space, ie between the inner cover plate and the housing, sometimes a very high pressure is formed by the split water flow, which corresponds essentially to the upper water pressure, and tries to move the impeller in the axial direction, which causes it to large axial loads Storage and relatively large axial displacements of the impeller comes. In both resulting wheel side chambers, between the housing and the outer or inner cover plate, creates a split water flow whereby a certain proportion of the medium does not flow through the impeller and thus there is a loss of efficiency and a loss of power. In addition, in both sides of the wheel creates a very rapidly rotating water disc due to the resulting friction of the rotation of the shaft counteracts and thus unfolds a braking effect, which in turn further reduces the efficiency.
For these reasons, it is desirable to provide a nearly completely tight seal between the impeller and the housing.

Such a seal is e.g. from WO 02/23038 A1, in which essentially two types of seals, which consist of special sealing rings mounted on two hydrostatic bearings, are disclosed:

A first type in which the sealing ring is rotatably mounted relative to the housing and the impeller and a hydrostatic bearing is supplied by flexible lines leading from the turbine housing to the sealing ring with the storage medium. Due to the prevailing operating conditions described above, such flexible lines are of course a certain vulnerability, so must be built accordingly robust and maintenance times must be shortened accordingly to prevent by regular maintenance breakage due to wear of the flexible cables, resulting in failure of the seal and can lead to significant damage. In addition, the installation of such a ring with a number of flexible lines is relatively expensive. For these reasons, such a design of a seal is rejected by both the manufacturers and the operators.

The second type relates to sealing rings, which are freely rotating "mounted" with respect to the impeller and the housing, wherein the storage medium for the hydrostatic bearing is fed through bores in the turbine housing and through connecting holes in the sealing ring itself. WO 02/23038 now shows in particular two variants of such a sealing ring.

In the first variant (according to FIG. 3 of WO 02/23038), a number of supply lines are provided, the openings of these supply lines facing into the bearing surfaces of the mouths of the connecting holes in the sealing ring. In practice, this ring shows unsatisfactory functionality. If the sealing ring is radially fixed to the housing, then the sealing ring can be made difficult to lift in the radial direction, the entire storage medium which is pressed in via the supply line is namely conducted via the connecting bore to the second bearing and leads to a strong lifting in the axial direction , The sealing ring would thus grind on the housing, which leads to damage and can result in destruction. On the other hand, in the case of the ring being installed with some radial play, it would center in operation and rise radially and axially, but it would not be in a preferred position due to the lack of balance of forces, since it would be unstable and he would be just as little adjustable radially. If one tries to change namely the radial position by changing the volume flow, only the axial position would change, since a changed volume flow would in turn be forwarded via the connecting bore directly to the axial bearing. Such a sealing ring would therefore be less practical in practice.

In the second variant (according to FIGS. 4 and 5 of WO 02/23038), at least two rows of supply lines are now provided, which are arranged at a distance from one another and via which storage medium is conveyed independently of one another into the hydrostatic bearings. The mouth of only one of these two supply lines stands opposite the mouths of the connecting hole in the sealing ring.
The advantage of this ring that both the radial and the axial bearing can be largely controlled and controlled separately and a stable operating position can be achieved in this variant has the disadvantage that for stabilization and controllability of both bearings two rows of supply lines are needed , which must be supplied and controlled independently of each other, ie at least two sets of supply pumps, some of considerable power, including the associated control or additional hydraulic components, such as throttles, filters, etc., are required, so that the performance gain an effective seal is partially or even completely consumed by the required pumping capacity or by throttling losses. In addition, the production of such a seal manufacturing technology is considerably more expensive, since of course twice the number of holes and lines are needed.

The subject invention has now taken on the task of eliminating the above-mentioned disadvantages and to provide an effective and reliable seal of the type mentioned, which requires little resources, can be easily implemented and operated, and has a long life.

This object is achieved by the invention by at least one further, second bearing element of the first hydrostatic bearing is arranged at a distance from a first bearing element of a first hydrostatic bearing, which is connected via a hydraulic resistor to the first bearing element, wherein the supply line for this Bearing opens exclusively in the area of the first bearing element in the bearing surface.
Such a sealing body reduces the required number of supply lines and thus reduces the manufacturing complexity, as well as the number of required supply units and elements.
Although only a single supply line is provided, one succeeds To set stable sealing body by means of a hydraulic resistance on a desired axial and radial bearing gap, resulting in a stable operating position. Both bearing gaps can be varied by the volume flow and stand in a substantially fixed relationship to each other, ie the sealing body is completely controllable with only one supply line in both directions. Since the sealing body rises substantially simultaneously and uniformly in the axial and in the radial direction, it is ensured that the sealing body does not rise in only one direction, which considerably improves operational reliability.

For the method for operating a seal according to the invention, the object is achieved in that the hydraulic machine is switched on only after the predetermined bearing gaps have been set. Characterized friction or Mischreibzustände and associated wear, damage or even destruction of the sealing body during startup of the machine are effectively prevented. The life of such a sealing body is thus significantly improved.

During the commissioning of the seal, so in one of the initial power-on, but it may be advantageous to bring the sealing body controlled in a mixed friction state, so that a bearing image can grind into the bearing surfaces of the hydraulic bearings. In order for certain manufacturing tolerances of the sealing body or the bearing surfaces are compensated and the operation, and the life of the seal can be improved. After grinding the bearing image of the sealing body is naturally raised to the predetermined bearing gaps and operated normally.
Since the sealing body is typically made of a softer material than the associated bearing surfaces on the housing or impeller or vice versa, such a bearing image can be achieved very easily and controlled.

The sealing body can be very easily manufactured and operated when the two hydrostatic bearings are connected to each other by means of a hydraulic connection. Thus, it is sufficient to supply only a single hydrostatic bearing with a storage medium, whereby the second is automatically supplied with.

A very advantageous pressure distribution, which leads to a safe lifting of the sealing body in both directions, adjusts by the width of the bearing element of the first hydrostatic bearing relative to the total width of this bearing smaller than the width of the bearing element of the second hydrostatic bearing based on the total width this warehouse is chosen. For the safe operation of the sealing body and achieving sufficient stability, it is particularly advantageous if the distance between the two bearing elements of the hydrostatic bearing in which the supply line opens is smaller than a maximum distance, which essentially results from the geometric dimensions of the sealing body. By complying with this geometric specification gives a very effective and reliable sealing body.

A particularly simple sealing body results in the form of a sealing ring. Such a ring can be made very easily and inexpensively.

The life of the sealing body is significantly increased when the sealing body is mounted on the hydrostatic bearings flying, because then a solid friction between the sealing surface and sealing body is excluded in all operating points of the hydraulic machine.

The seal according to the invention is advantageously used for sealing an impeller and a housing of the hydraulic machine, in particular a turbomachine, with which the wheel side spaces can be effectively sealed and, with a corresponding arrangement, e.g. in the peripheral area of the impeller, with the exception of the storage medium, are not filled with the operating medium of the hydraulic machine. The formation of the above-mentioned negative effects are thereby prevented.

Especially advantageous is the application of the seal for a turbine, in particular a Francis turbine or a pump turbine, or a pump.

The bearing element can be easily and inexpensively formed as a, optionally interrupted in sections, annular groove over the circumference, which is also very easy to produce. Likewise constructive and manufacturing technology are simple bores as a hydraulic connection in the sealing body and as a supply line in the housing of the hydraulic machine.

The properties of the sealing body and thus the seal itself can be further improved by the arrangement of a third bearing element, or more bearing elements in the bearing surfaces of the hydrostatic bearing. The additional bearing element creates a broader pressure distribution, which is easier to control and with which the torque balance on the sealing body can be adjusted more easily.

The sealing body is carried out in an advantageous manner such that the central bearing element is made wider than the other bearing elements. Another favorable geometric specification results from the special choice of the distance between the outer edges of the two outer of the plurality of bearing elements based on the width of this hydrostatic bearing, so that it is smaller than the width of the bearing element of the other hydrostatic bearing based on the width of this hydrostatic bearing.
Likewise, it is favorable for given geometric dimensions of the sealing body, such as height and width of the sealing ring, arrangement and width of the bearing elements, in particular the grooves, grooves, etc., the distance between the first and the second bearing element of the first hydrostatic bearing smaller one to select a predetermined maximum distance.

The hydrostatic bearings are very advantageously supplied with a constant volume flow of the storage medium. Thus, the sealing body independently and controllably to changes in external conditions, such as temperature changes of the medium and an associated change in length of the sealing body, vibrations of the housing or the impeller, manufacturing tolerances, tilting of the sealing ring, etc., react, since the flow, in addition to geometric dimensions, significantly responsible for the pressure distribution. The sealing body is thus self-regulating, so it compensates for external disturbances independently.

A simple supply of the hydrostatic bearing can be ensured by at least one pump. As a possible alternative to pumps, the upper water, which naturally has a high hydrostatic pressure, could also be used, and then at least one throttle, e.g. designed as a flow control valve, should be provided in order to specify a certain substantially constant volume flow can.

The sealing body or the seal can be operated extremely favorably and with low losses if the power loss caused by the sealing ring is minimized by a suitable geometry of the sealing body. Such a seal thus has a minimal power loss, whereby the overall efficiency of the turbine, due to the prevention of the formation of fission water flows through the seal, can be considerably increased.

The bearing effect of the hydrostatic bearing can be considerably increased in part if at least one hydrodynamic bearing element, such as a lubricating pocket, is additionally provided in at least one of the bearing surfaces. In addition to the conventional hydrostatic bearing action, a hydrodynamic bearing effect is additionally added, which at the prevailing speeds can make up a considerable proportion of the total bearing effect.

In case of failure of the hydraulic bearing supplying flow rate, the hydraulic machine should preferably be turned off to any damage to prevent the seal or on the sealing body. The reliability is increased if, in the case of a failure, at least for a certain period of time, preferably until the hydraulic machine has come to a standstill, an emergency supply of hydraulic bearings, eg by means of an air chamber is ensured. By such emergency supply possible damage to the sealing ring can be avoided.

The efficiency can be further improved by turning off a number of supply sources after the hydraulic machine has started up. It should of course be ensured that the remaining supply is sufficient to keep the sealing body in all operating states in the flying state, without causing mixed friction phases.

Substantially constant bearing gaps can be ensured in a very simple manner, if natural changes in the seal body geometry, such as e.g. the swelling of the sealing body in the medium can be compensated by varying the volume flow supplied.

• Fig. 1 zeigt einen Querschnitt einer typischen Francisturbine,
• Fig. 2 eine Detailansicht des Dichtbereiches zwischen Laufrad und Gehäuse mit einem erfindungsgemäßen Dichtkörper,
• Fig. 3 eine weitere spezielle Ausführungsform eines Dichtkörpers und
• Fig. 4 ein geometrische Verhältnisse des Dichtkörpers darstellendes Diagramm.

The subject invention will now be described by way of example, with reference to FIGS. 1-4, which show non-limiting specific embodiments.

• 1 shows a cross section of a typical Francis turbine,
• 2 is a detail view of the sealing area between the impeller and housing with a sealing body according to the invention,
• Fig. 3 shows a further specific embodiment of a sealing body and
• Fig. 4 is a geometric relationships of the sealing body darstellendes diagram.

In advance of the actual description, some terms are defined and explained in more detail below.
Often the terms bearing elements, such as grooves, and bearing surface are used, which in the context of this application each describe a ring-shaped or cylindrical structure. A bearing element or bearing surface can have any widths and depths or heights and can be interrupted in the circumferential direction continuously or else in sections at one or more points. Of course, a bearing element can have any desired cross-sectional shape, eg also a triangular groove, and does not necessarily have to be designed as a groove.
A hydraulic bearing always consists of mutually facing bearing surfaces, wherein in at least one of the bearing surfaces at least one bearing element, such as a groove, grooves or the like, is arranged. If a plurality of bearing elements are now arranged over the circumference, because a bearing element is interrupted in sections, for example, as described above, then so Consequently, one would also have to speak of several hydrostatic bearings distributed over the circumference. For the sake of simplicity, however, in the application, even in such cases, only one hydrostatic bearing is used. For the purposes of this application, hydraulic connection or connection bore means at least one cavity with at least two open ends, wherein this cavity can be flowed through by any medium from one end to the other end in any desired way.
When talking about a supply line, it should be noted that a number of such similar or similar supply lines, ie a number of supply lines, can be arranged in the circumferential direction. Of course, the same applies to a connection hole. However, in order not to complicate the description, it is usually spoken only of a supply line or a connection bore, which of course also includes a number of supply lines or connection holes, if necessary.

For the sake of simplicity, the seal according to the invention will be described only with reference to a turbine, in particular a Francis turbine, but of course also in all other hydraulic machines with relatively movable parts, such as an impeller running in a machine housing, such as for pumps or pump turbines, is equivalent applicable.

1 now shows a turbine 1, here a Francis turbine, with an impeller 2, which runs in a turbine housing 12. The impeller 2 has a number of turbine blades 3 which are bounded by an inner 11 and outer cover plate 10. The impeller 2 is connected by means of a hub cap 9 and possibly by means of further attachment means, e.g. Bolts or screws, secured at one end of the shaft 8 with respect to the shaft 8 rotationally. The shaft 8 is rotatably supported by means not shown shaft bearings and drives in a known manner, for example, a generator, also not shown, for generating electrical energy, which is preferably arranged at the other end of the shaft 8.

The inflow of the liquid medium, usually water, from an upper water, such as a higher-lying water reservoir, in most cases via a not shown here, well-known volute casing. Between spiral housing and impeller 2 is a diffuser 4, consisting of a number of vanes 5, which are rotatable in this example by means of an adjusting device 6. The adjustable vanes 5 serve to regulate the power of the turbine 1 by varying both the volume flow through the turbines 1 and the impeller inlet swirl. In addition, between the volute casing and guide vanes 5 could also be arranged in a known manner supporting blades.

The outflow of water takes place, as shown in Fig. 1, via a directly adjacent to the turbine 1 suction pipe 13, which opens into an unillustrated underwater. This results in a marked by the arrow main water flow F from the volute casing via the nozzle 4 and the impeller 2 to the suction pipe thirteenth
In addition to the main water flow F is formed in conventional seals 14 and a gap water flow through the wheel side spaces between the turbine housing 12 and outer 10 and inner cover plate 11 from. The split water of the radial wheel side space is discharged, for example, by means of a line 7 via a throttle and passed into the suction pipe 13. In addition, as indicated in Fig. 1, relief bores are often provided in the inner shroud through which the radial wheel side space is connected to the main water flow F. As a result of the sealing according to the invention, as described below, these fissured water streams are now prevented, so that the entire inflowing water flows through the impeller 2 and its flow energy, without any gap losses, can be fully utilized. In addition, the friction in the wheel-side space is reduced or even minimized, since with such a seal in the wheel side space no rotating water disc longer forms, but this space, with the exception of the storage water, is filled by air. Furthermore, this also greatly reduces the axial thrust acting on the shaft 8 and on the shaft bearing.

FIG. 2 shows a detailed view of an exemplary inventive seal of an impeller 2 of a turbine 1 between turbine housing 12 and inner cover disk 11 by means of a sealing body 20 designed as a sealing ring. In this case, the turbine housing 12 has a shoulder on which a radial bearing surface 24 is arranged. Likewise, an axial bearing surface 23 is disposed on the inner cover plate 11. These bearing surfaces 23, 24 may be separate components, which are subsequently applied to the required location, for example by welding, screws, etc., or can of course also be incorporated in the corresponding component, eg a surface-ground section on the inner cover disk 11.
The orientations "axial" and "radial" relate to the effective directions of the hydrostatic bearings and are mainly introduced for the easier differentiation of the two hydrostatic bearings 21, 22.
The radial 24 and the axial bearing surface 23 on the turbine housing 12 and on the inner cover plate 11 is each assigned a radial 24 or axial bearing surface 23 on the sealing ring 20, each forming a part of a hydrostatic bearing 21, 22 in the axial and radial directions ,
In the embodiment according to FIG. 2, a storage medium, for example water, is fed into the radial hydrostatic bearing 22 via the turbine housing 12 by means of a supply line 28. The supply line 28 is formed here from holes that have more Lines indirectly or directly to a supply source, not shown, such as a pump and / or the upper water, possibly via auxiliary equipment such as filters, cyclones, etc., is connected. Of course, a plurality of supply lines 28 may be distributed over the circumference, wherein a favorable arrangement for a sufficient supply, for example three supply lines 28, which are each arranged offset by an angle of 120 °, may be provided. Of course, any other arrangement is conceivable.
The radial bearing 22 now has two bearing elements in the form of grooves 25, 26, wherein a groove 25 in the sealing ring 20 in the region of the mouth of the supply line 28 and the second groove 26 is also disposed in the sealing ring 20 at a distance from the first groove 25. This second groove 26 is now connected via one or more connecting bore (s) 29 with a groove 27 of the axial bearing 21 arranged in the sealing ring 20. The two grooves 25, 26 of the radial bearing 22 are now arranged such that the supply line 28 opens in all operating positions of the sealing ring 20 neither completely nor partially into the second groove 26.
It should be noted in particular also that the bearing elements, here grooves 25, 26 and 27, also equivalent in the axial or radial bearing surface 23, 24 of the turbine housing 12 or the impeller 2, as here in the inner cover plate 11, could be arranged , It would also be possible to provide bearing elements both in the sealing body and in the turbine housing 12 or at any point of the impeller 2.
Thus, both hydrostatic bearings 21, 22 are supplied by a single supply line 28 or series of supply lines 28 with storage medium. The storage medium is supplied in the radial bearing 22 and flows through the connecting holes 29 in the axial bearing 21. In order to ensure adequate supply of the axial bearing 21, advantageously over the circumference of the sealing ring 20 a plurality of connecting holes 29 are provided, depending on Circumference eg a hole every 3 to 8 centimeters.
The groove 27 of the axial bearing 21 could equivalently also be designed so that in each case a narrower groove is disposed in the region of the outer and inner diameter of the sealing ring 20, which is connected via a connecting bore 29 with the radial bearing 22 and is supplied.
The supply line 28 could of course also open in the axial bearing 21, in which case the arrangement of the grooves 25, 26 and 27 would also mirrored accordingly on the diagonal of the sealing ring.

In order to describe the function of the sealing ring 20, the resulting pressure distributions in the axial and radial bearings 21, 22 are additionally shown in FIG. The storage medium is supplied via the supply line 28 with a constant volume flow Q in the axial bearing 21 as described above. The volume flow Q of the storage medium is divided in the radial bearing 22 into two streams. A current flows down and finally ends with the pressure p ₀ in the axial wheel side space. The greater part of the volume flow Q flows upward to the second groove 26 and flows through the connecting hole 29 in the radial bearing 21 and opens partially with the prevailing at the impeller inlet pressure p ₁ in the storage room 31st
The volume flow Q causes the illustrated pressure distribution with a maximum pressure p ₃ in the groove 25, in which the supply line 28 opens, which lifts the sealing ring 20 in the radial direction. Radial lifting in this context, of course, means that the sealing ring 20 expands, this expansion of both the upper water pressure p ₁ , and according to the theory of elasticity counteract the elastic restoring forces. The maximum pressure p ₃ must therefore be large enough to effect such expansion of the sealing ring 20 to the desired bearing gap, for example typically 50-100 μm. In the second groove 26 is due to the geometry of lower pressure p ₂ , which simultaneously acts through the connecting hole 28 in the groove 27 of the axial bearing 21. This pressure p ₂ must be sufficiently high to lift the sealing ring 20 in the axial direction, which can be achieved by placing the two grooves 25, 26 of the radial bearing 22 very strongly asymmetrically and very close together, as in FIG Fig. 2 shown.
If the grooves 25, 26 were too far apart, the pressure drop between the grooves 25, 26 would be too great and the required lift-off pressure p ₂ would not be achieved. That is, the pressure p ₂ for the example of FIG. 2 by the geometry of the axial bearing of the sealing ring 20, so substantially width and position of the grooves of the corresponding bearing 22, outer dimensions of the sealing ring 20, and any recesses 30, is set. If the volume flow Q were then to be increased further, the pressure p _{2 would} nevertheless remain substantially the same and the sealing ring 20 would only lift off further in the axial direction.

By using basic hydraulic principles for any desired geometry of the sealing ring 20 or any arrangement of the grooves 25, 26, 27, a maximum distance f _max between the two grooves 25, 26 can be specified, which depends exclusively on the geometry and complied with must be to bring the sealing ring 20 in both directions to take off. Determining the maximum distance f _max represents a standard task for a corresponding person skilled in the art. In FIG. 4 (which relates here to the geometry of FIG. 2), an ascertained profile of the maximum distance f _{max is} shown by way of example. In this example, the outer dimensions of the sealing ring 20, as well as the geometric dimensions of the axial hydrostatic bearing 21 and certain geometric dimensions of the radial hydrostatic bearing 22 are kept constant and only the distance d of the upper edge of the sealing ring 20 to the second groove 26 varies and the result in the form of a diagram in FIG. 4 shown. The sizes used in the diagram was related to the width B _{r of} the radial bearing 21 and therefore made dimensionless. The point shown in Fig. 4 shows the distance f according to the geometry of FIG. 2. It can be clearly seen that the sealing ring is in the stable range.
If other geometric parameters are varied, it is of course possible to obtain other shapes of the curve or area, for example, by varying two parameters. Of course, similar relationships can also be stated for other embodiments of a sealing ring 20, for example as described in FIG. 3.

The sealing ring 20 so now floats stable on two sliding films practically smoothly, so it is "flying" stored. In operation, the sealing ring 20 will rotate due to the free storage with approximately half the peripheral speed of the impeller 2, since it is not held against rotation. This results in a dynamic stability gain, since thus the limit peripheral speed or the flutter limit is set up. In addition, this also reduces the friction losses.
Due to the high stability of a hydrostatic bearing, the sealing ring 20 is able to compensate for vibrations of the impeller 2 and / or the turbine housing 12, and axial displacements of the impeller 2, without losing the sealing effect and without being in contact with the impeller 2 and / or to get the turbine housing 12. The sealing ring 20 suffers practically no wear, whereby the life of such a sealing ring 20 is very high. The fact that the sealing ring 20 can be built as a very slender, lighter ring, which has hardly any inertial forces, this effect is further enhanced.

The sealing ring 20 can be made very small in relation to the dimensions of the turbine 1, edge lengths of a few centimeters, e.g. 5cm or 8cm, with outside diameters of a few meters are quite sufficient, and it can be made of any material such as steel, bearing bronze, plastic (e.g., PE). Furthermore, the bearing surfaces 23, 24 could also be coated with a suitable layer, such as Teflon, bearing bronze, etc., in order to further improve the properties of the seal. Typically, the sealing ring 20 is made of a softer material such as the housing 12 or the impeller 2 of the hydraulic machine. As a result, he is on the one hand usually lighter and on the other hand, in extreme cases, the sealing ring 20 and not the impeller 2 or the housing 12 is damaged or even destroyed.

Since the sealing ring 20 can be very small in cross-section, but can act very high pressures, there is a risk of Verkrempelns the sealing ring 20. To compensate for the resulting carding moments, the sealing ring 20 should be designed torque-free, ie the sealing ring 20 should in Operation have no resulting moment. This can, as one can easily think, be achieved by: the sealing ring 20 is designed so that the resulting forces of the respective pressure distributions on the sides of the sealing ring 20, so the resulting forces of the upper water pressure p ₁ and the resulting pressure distributions in the hydrostatic bearings 21, 22, lie on a line of action. To achieve this, in addition to the entire geometry of the sealing ring 20, such as the dimensions of the grooves 25, 26, 27, the bearing gap widths, the outer dimensions, etc, among other things, the recess 30 is used.

The sealing ring 20 may, of course, be of any cross-section, e.g. an L-shaped cross section, wherein for manufacturing considerations, a square or rectangular shape is preferred.

In Fig. 3, another embodiment of an inventive sealing ring 20 is shown. This sealing ring 20 now has in the radial bearing 22 three grooves 25, 26, wherein in the region of the central groove 25, as described in Fig. 2, a supply line 28 opens, via which a volume flow Q of a storage medium is supplied. The two grooves 26 arranged laterally from this central groove 25 are each connected via connecting bores 29 to the two grooves 27 of the axial hydrostatic bearing 23. In this example, two grooves 27 are arranged, which unfolds the same effect as a continuous groove 27, as described in Fig. 2. As the width of the groove of the axial hydrostatic bearing 23 so the distance between the outer diameter of the left and inner diameter of the right groove 27 can be considered.
Each of the two outer grooves 26 of the radial bearing 22 is here connected to each of the grooves 27 of the axial bearing 23 via a system of connecting bores 29 which are arranged in a cross-sectional plane of the sealing ring 20. However, it is also conceivable to separate the connections and to arrange them in different cross-sectional planes of the sealing ring. In a cross-sectional plane, for example, the upper groove 26 would be connected to the right groove 27, in a next cross-sectional plane, the lower groove 26 with the left groove 27 and again in a next, a system of connecting holes 29, as shown in Fig. 3, arranged be. Of course, if required, any combination is possible.

Considering the pressure distributions of this sealing ring 20, it can be seen that the pressure distribution with respect to the embodiment of FIG. 2 in the axial hydrostatic bearing 21 remains substantially unchanged within the geometric conditions, while the pressure distribution in the radial hydrostatic bearing 22 changes considerably. This pressure distribution is now, caused by the third groove 26 wider and has lower pressure peaks, ie, that such a sealing ring 20 is operable with a lower supply pressure.

Of course, the three grooves 25, 26 of the radial hydrostatic bearing can be arranged substantially arbitrarily. For example, the two outer grooves 26 could have the same width and be arranged symmetrically about the central groove 25 or with respect to the sealing ring 20 itself. On the other hand, the arrangement of the three grooves 25, 26 could also be completely asymmetrical.
Likewise, it would be conceivable to provide more than three grooves, which could possibly be achieved even flatter pressure distribution.

In the examples according to FIGS. 2 and 3, the supply line 28 always opens into the radial hydrostatic bearing 22, whereas the axial hydrostatic bearing 21 is supplied by connecting bores 29. If necessary, this arrangement can of course also be executed vice versa.

In addition, until now, it has always been assumed that flat bearing surfaces 23, 24 are used. However, it is of course also conceivable that the bearing surfaces 23, 24 are not flat, e.g. concave ground or graded to execute, with the fundamental principle of the seal according to the invention does not change. With such non-planar bearing surfaces 23, 24, only the pressure distributions would change somewhat, but this is clearly apparent to a person skilled in the art.

The operation of the sealing ring 20 creates a certain power loss, for example by the required power of one or more supply pump (s), by hydraulic friction in the bearing gap, by underwater losses, ie storage medium that can not be passed through the impeller 2, etc., the should be kept as low as possible. Of course, part of this power loss can be recovered by passing part of the storage medium into the main water flow F and converting it into energy in the impeller 2. Nevertheless, it is desirable to minimize the power dissipation of the sealing ring 20. For this purpose, the geometry of the sealing ring 20, so width, height, position and dimensions of the grooves 25, 26, 27 and bearing surfaces 23, 24, dimensions and position of the supply lines 28 and the connecting holes 29, etc., adjusted to the resulting power loss minimize. This can be done, for example, by suitable mathematical, eg numerical, calculations on the basis of mathematical, physical models of the sealing ring, in which an appropriately formulated optimization problem is solved. With conventional computers and corresponding software, such an optimization problem can be solved. Of course, the geometry and / or the operating characteristics, eg design points, outputs, pressures, etc., of the hydraulic machine can also be included in these calculations.

To improve the bearing effect and the stability, one or more of the bearing surfaces 23, 24 could also be provided with well-known hydrodynamic lubrication pockets.

Basically, a number of supply lines 28 will be merged into a large manifold, which will then be fed from a storage medium source, such as a storage medium. a pump, to be supplied with storage medium. Of course, the number of storage medium sources and manifolds can be freely selected as needed.

A seal according to the invention with a sealing ring 20 can of course be provided at any suitable location and is not limited to the embodiments of FIGS. 2 and 3. For example, the sealing ring 20 could also be arranged between the end face of the impeller 2 or inner cover disk 11 and the turbine housing 12. Similarly, it is conceivable to provide such a seal at a suitable location between the outer cover plate 10 and the machine housing 12.

In addition, the arrangement of the grooves 25, 26, 27, and the connecting holes 29 and supply line 28 in Figs. 2 and 3 is merely exemplary. Rather, this arrangement can be chosen arbitrarily within the scope of the invention. For example, could the groove 26, which is connected via the connecting hole 29 with the groove 27 of the other hydrostatic bearing 21, equivalently in the vicinity of the recess 30, so in Fig. 2 below the groove 25, are arranged. The entire assembly could also be mirrored on the diagonal of the sealing ring. All possible and conceivable variants are of course included in this application.

The seal described above is a largely tight seal. The entire amount of water flowing through the impeller and can be converted into rotational energy. The splitting water losses are reduced exclusively to the emerging storage medium, so they are very low and can be recovered in part by introducing the cracked water into the main water flow F again.

In all phases of the operation should be avoided as far as possible that the sealing ring 20 comes into contact with the bearing surfaces 23, 24 or set mixed friction conditions in the hydraulic bearings 21, 22, since then the sealing ring 20 can be easily damaged or even destroyed. When starting the turbine 1, the sealing ring 20 should therefore already be lifted, ie the desired bearing gaps should already be reached. This can be easily achieved by first switching on the supply of the hydraulic bearings 21, 22 and only then the turbine 1 is turned on.
In case of failure of the supply of the hydrostatic bearings 21, 22 could be, for example provide an emergency supply, such as a wind turbine, to avoid damage to the sealing ring 20 of the hydraulic machine, which would entail extensive maintenance.

However, during the initial start-up of the sealing ring 20, it may be desirable to set a controlled mixed friction state in the hydraulic bearings 21, 22 so that a bearing image can grind into the bearing surfaces 23, 24, whereby certain production inaccuracies can be compensated. Since the bearing gaps are in the hundred μm range or below, of course, appropriate caution is appropriate.

In the description above, water is described as the storage medium for the sake of simplicity. Of course, the storage medium, especially in pumps, can also be any other suitable medium, e.g. an oil, his.

For the sake of clarity throughout the application is always spoken of grooves, grooves or the like as bearing elements. However, it is quite conceivable not to make one or more bearing elements so clear. Of course, any gap flow, even between groove-free smooth surfaces, has some hydraulic resistance, so that a hydrostatic bearing can be used even without pronounced bearing elements, e.g. only with flat surfaces, would work. Surface roughness would also result in further influencing of the hydraulic resistance, e.g. could the bearing surfaces 23, 24 are ground to form "bearing elements" different.

Claims (34)

1. An apparatus for sealing a gap between two mutually mobile parts of a hydraulic machine with at least one sealing element which is mounted with respect to the two mobile parts by means of at least one hydrostatic bearing in each case, each of the hydrostatic bearings comprising mutually facing bearing surfaces and at least one bearing surface having at least one bearing element, such as a groove, flute or the like, which can be supplied with a hydraulic bearing medium via at least one supply line, characterized in that, at a distance from the first bearing element of a first hydrostatic bearing, at least one further, second bearing element of the first hydrostatic bearing is arranged, which is connected to the first bearing element via a hydraulic resistance, the supply line for this bearing opening only into the bearing surface in the region of the first bearing element.
2. The apparatus as claimed in claim 1, characterized in that the second bearing element of the first hydrostatic bearing is connected to a bearing element of the bearing surface of the second hydrostatic bearing by means of a hydraulic connection.
3. The apparatus as claimed in claim 2, characterized in that the first bearing element of the first hydrostatic bearing has no direct hydraulic connection to the second hydrostatic bearing.
4. The apparatus as claimed in claim 1, 2 or 3, characterized in that the second bearing element of the first hydrostatic bearing is not directly assigned any opening of a supply line.
5. The apparatus as claimed in one of claims 1 to 4, characterized in that the width of the second bearing element of the first hydrostatic bearing, as based on the width of this first hydrostatic bearing, is smaller than the width of the bearing element or elements of the second hydrostatic bearing, as based on the width of this hydrostatic bearing.
6. The apparatus as claimed in one of claims 1 to 5, characterized in that the sealing element is formed as a sealing ring.
7. The apparatus as claimed in one of claims 1 to 6, characterized in that the sealing element is arranged mounted in a floating manner on the hydrostatic bearings.
8. The apparatus as claimed in one of claims 1 to 7, characterized in that the mobile parts of the hydraulic machine are an impeller and a housing of the hydraulic machine, in particular of a turbo machine.
9. The apparatus as claimed in one of claims 1 to 8, characterized in that the hydraulic machine is a turbine, in particular a Francis turbine or a pump turbine.
10. The apparatus as claimed in one of claims 1 to 9, characterized in that the hydraulic machine is a pump.
11. The apparatus as claimed in one of claims 1 to 10, characterized in that the bearing element is formed as an annular groove which may be interrupted in sections over the circumference.
12. The apparatus as claimed in one of claims 2 to 11, characterized in that the hydraulic connection in the sealing element is at least partly formed as a bore.
13. The apparatus as claimed in one of claims 1 to 12, characterized in that the supply line is at least partly formed as a bore in the housing of the hydraulic machine.
14. The apparatus as claimed in one of claims 1 to 13, characterized in that, given predefined geometric dimensions of the sealing element, such as the height and width of the sealing ring, arrangement and width of the bearing elements, in particular of the grooves, flutes, etc., the distance between the first and the second bearing elements of the first hydrostatic bearing is smaller than a predetermined maximum distance.
15. The apparatus as claimed in one of claims 1 to 14, characterized in that three or more bearing elements spaced apart from one another are provided in the bearing surfaces of the first hydrostatic bearing, the at least one supply line opening into the central bearing element and, of the remaining bearing elements of the first hydrostatic bearing, at least two bearing elements being connected to at least one bearing element of the second hydrostatic bearing via at least one hydraulic connection in each case.
16. The apparatus as claimed in claim 15, characterized in that the central bearing element is designed to be broader than the remaining bearing elements of the first bearing.
17. The apparatus as claimed in claim 15 or 16, characterized in that the first hydrostatic bearing has a plurality of bearing elements, for example three bearing elements, and the second hydrostatic bearing has one bearing element, the distance between the outer edges of the two outer of the bearing elements, as based on the width of this hydrostatic bearing, being smaller than the width of the bearing element of the second hydrostatic bearing, as based on the width of this hydrostatic bearing.
18. The apparatus as claimed in one of claims 1 to 17, characterized in that the hydrostatic bearings can be supplied with a substantially constant volume flow of the bearing medium through the supply line.
19. The apparatus as claimed in one of claims 1 to 18, characterized in that the supply line or a series of supply lines is connected to at least one pump.
20. The apparatus as claimed in one of claims 1 to 19, characterized in that the supply line or a series of supply lines is connected to the headwater of a hydraulic machine having the seal.
21. The apparatus as claimed in claim 19 or 20, characterized in that at least one restrictor is provided upstream of the opening of the supply line or the series of supply lines into the hydrostatic bearing.
22. The apparatus as claimed in claim 21, characterized in that the restrictor is designed as a flow regulating valve.
23. The apparatus as claimed in one of claims 1 to 22, characterized in that the geometry of the sealing element, for example width, height, position and dimensions of the bearing elements and bearing surfaces, dimensions of the supply line and of the hydraulic connections, etc., can be predefined in such a way that the power loss of the seal substantially assumes a minimum.
24. The apparatus as claimed in one of claims 1 to 23, characterized in that at least one hydrodynamic bearing element, preferably a lubrication pocket, is provided in at least one of the bearing surfaces.
25. A method of operating a seal for a gap between two mutually mobile parts of a hydraulic machine, comprising at least one sealing element which is mounted with respect to the two mobile parts by means of at least one hydrostatic bearing in each case, characterized in that, before the hydraulic machine is switched on, at least one first hydrostatic bearing is supplied via a supply source with a volume flow, preferably a substantially constant volume flow, of a hydraulic medium, the sealing element being lifted with respect to the sealing surfaces of the hydrostatic bearings and predefined bearing gaps being established in a substantially stable manner, and in that the hydraulic machine is then switched on.
26. The method as claimed in claim 25, characterized in that, in the event of failure of the volume flow supplying the hydraulic bearings, the hydraulic machine is switched off.
27. The method as claimed in claim 25 or 26, characterized in that, in the event of failure of the bearing supply, the hydraulic bearings are supplied from an emergency supply source, such as an air reservoir or an emergency supply reservoir, at least over a certain time period.
28. The method as claimed in one of claims 25 to 27, characterized in that, after the hydraulic machine has been run up, the volume flow is substantially reduced to a minimum, for example by a number of the supply sources being switched off.
29. The method as claimed in one of claims 25 to 28, characterized in that natural changes in the geometry of the sealing element, such as those caused by temperature influences, centrifugal force effects, the swelling of the sealing element in the medium, etc., are compensated for by varying the volume flow supplied in such a way that the bearing gaps remain substantially constant.
30. A method of commissioning a seal for a gap between two mutually mobile parts of a hydraulic machine with at least one sealing element which is mounted with respect to the two parts by means of at least one hydrostatic bearing in each case, characterized in that, before the hydraulic machine is switched on, at least one hydrostatic bearing is supplied with a substantially constant volume flow in such a way that the sealing element is lifted with respect to the sealing surfaces of the hydrostatic bearings and predefined bearing gaps are established, in that the hydraulic machine is then switched on, and in that the volume flow is then reduced in a controlled manner until the bearing surfaces of the sealing element and the associated bearing surfaces on the hydraulic machine change into a frictional state, preferably a mixed friction state and begin to rub, so that a bearing pattern is ground into the bearing surfaces.
31. The method as claimed in claim 30, characterized in that, after and/or during the grinding of a bearing pattern, the volume flow is increased, so that the bearing gaps predefined for the operation are substantially established.
32. A method of designing a seal for a gap between two mutually mobile parts of a hydraulic machine with at least one sealing element which is mounted with respect to the two parts by means of at least one hydrostatic bearing in each case, characterized in that a power loss of the seal is predefined and the geometry of the seal, for example width, height, position and dimensions of the bearing elements and bearing surfaces of the sealing element, dimensions of the supply line and of the hydraulic connections, etc., are calculated by using mathematical, physical models of the seal while taking account of the predefined power loss and/or of the geometry and/or of the operating characteristics of the hydraulic machine.
33. The method as claimed in claim 32, characterized in that the geometry is optimized with respect to the energy consumption, the power loss of the seal is therefore substantially minimized.
34. The method as claimed in claim 32 or 33, characterized in that the calculations are carried out in a computer.

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